Geared transmissions

ABSTRACT

A gear assembly ( 10 ) for transmitting torque between an input rotatable member ( 12 ) and an output rotatable member e.g. differential unit ( 24 ) comprises an input gear ( 14 ) rotatable with the input member ( 12 ), an output gear ( 20 ) rotatable with the output member ( 24 ), and intermediate gears ( 16   a   , 16   b   , 16   c   , 16   d ) held simultaneously in mesh with the input gear ( 14 ) and output gear ( 20 ). Various arrangements are disclosed for sharing the transmitted torque evenly between the paths provided by gears ( 16   a   , 16   b   , 16   c   , 16   d ). The transmission may provide a reduction ratio and is connected to a low reduction ratio differential gearbox of small size and slim profile.

This application claims priority to Great Britain Patent Application No.0015115.9 filed on Jun. 20, 2000 and International Application No.PCT/GB01/02667 filed on Jun. 15, 2001 and published as InternationalPublication No. WO 01/98100 A1 on Dec. 27, 2001, the entire contents ofwhich are hereby incorporated by reference.

This invention relates to transfer gearing for transmitting powerbetween first and second members rotating about parallel axes inautomotive transmissions, for example in which power from the outputshaft of a main gearbox must be transmitted to a pair of propellershafts respectively extending fore and aft of the vehicle. Thesepropeller shafts are usually interconnected by differential gearing, andextend generally parallel to the main variable ratio gearbox outputshaft. Typically, power is transmitted between a gear on the maingearbox output shaft and a similarly sized differential gear, to give anapproximate 1:1 gear ratio. There are other instances where power mustbe transmitted from a first shaft to a second, generally parallel shaftin an automotive transmission.

In heavy vehicles such as trucks, this power transmission is often viaan intermediate gear supported on a layshaft of a transfer gearbox. Thediameter and/or lateral offset of the intermediate gear may be selectedto provide the required centre spacing between the parallel shafts.Where a large centre spacing is required, the intermediate gear diametermust be made correspondingly large. As the maximum transmissible torquedepends mainly upon the load per unit tooth width, the intermediate gearand the co-operating gears must each be made relatively wide fortransmission of high torque loads.

High torque loads and large centre spacings thus lead to a large andheavy gear train between the main and transfer gearboxes. Large diametergears have a higher pitch line velocity and hence tend to be noisier inoperation than smaller gears.

For lighter 4×4 “sports utility” vehicles, such transfer boxes virtuallyall use chain drives for power transmission between the parallel shafts.The driveline must operate at high RPM and because of the necessary geardiameter/operating centre distances a conventional single power pathgeared drive as used in trucks, for example, would involve unacceptablylarge gears and pitch line velocities. At the speeds concerned, chaindrive has proven quieter than even precision ground gears and allows amore compact casing. However chain drives suffer from overheatingproblems and excessive centrifugal loadings if used at speeds aboveabout 6000 RPM. Their power transmission capacity is therefore limited.

Theoretically, an alternative design approach would be to provide a pairof intermediate gears offset to either side of the plane containing theparallel shaft axes, these intermediate gears each meshingsimultaneously with corresponding gears on the parallel shafts (e.g. inthe main and differential gearboxes respectively). Assuming that allfour such gears are perfectly concentric, with perfect tooth pitches andprofiles, supported on shafts perfectly spaced relative to one another,journalled in perfect, play-free bearings, the whole being made fromperfectly inelastic materials, the transmitted torque will be sharedequally between the two intermediate gears, to provide parallel torquetransmission paths. The intermediate gears and the co-operating gears inthe main and differential gearboxes could thus theoretically be madecorrespondingly smaller and lighter.

However, commercially manufactured gearboxes are not perfect. Thekinematic forces acting on the intermediate gears coupled with theelasticity of the materials of the gear assemblies means that in realityone of the intermediate gears tends to be forced inwards towards theplane of the input/output shafts, whilst the other intermediate geartends to be forced outwards away from that plane. The intermediate gearforced inwards experiences a higher torque than the intermediate gearforced outwards.

Furthermore, dimensional inaccuracies in the various gearbox componentsmeans that in reality one of the intermediate gears will, when torque isapplied, assume flank-to-flank drive contact with each of the twoadjacent gears, whilst at that instant the opposite intermediate gearhas not established drive contact. Thus at that time only one torquetransmission path is effective. As the torque load increases, providedthat dimensional inaccuracies are within acceptable limits, gearboxcomponents will deform under load until mutual drive contact isestablished between all adjacent gears. However, torque sharing betweenthe two transmission paths will be unequal, with the degree ofinequality corresponding to the size of the dimensional inaccuracies.The proportion of the torque transmitted through each path may varythroughout the rotation cycle of the gearbox assembly, as thedimensional inaccuracies of each gear may vary cyclically.

Studies by NASA on helicopter gearboxes (see paper by Timothy L; Krantz“A Method to Analyze and Optimize the Load Sharing of Split-PathTransmissions”, published in Design Engineering, vol. 88, PowerTransmission and Gearing Conference ASME 1996, at pages 227-242) haveshown that satisfactory torque sharing between two parallel transmissionpaths can be achieved if tooth flank position errors are controlled toless than 0.0005 radian. Under these conditions, the inequality oftorque transmission might not vary beyond, say, 60:40, leading toworthwhile savings in gearbox size and weight. Such dimensional accuracyis achievable, certainly in aerospace and similar specialistapplications where high manufacturing costs are not prohibitive. Howeverthe requirement for high dimensional accuracy means that such torquesharing arrangements are impractical for mass produced automotivegearboxes, where low cost is an important factor.

U.S. Pat. No. 6,035,956 discloses an axle for low platform town buses inwhich hub reduction gear trains are connected, one on each side, betweenthe axle differential and respective offset stub axles carrying the busroad wheels. Each transfer gear train comprises a pair of intermediategears providing parallel power transmission paths. An input gear fixedto a respective output shaft of the differential meshes with bothintermediate gears simultaneously and is vertically movable so as toshare torque evenly between the power transmission paths.

U.S. Pat. No. 5,896,775 concerns high reduction final drive gearing fora powered scooter or wheelchair, in which an input shaft is connected todrive a pair of ground wheels through a pair of torque sharing piniongears. The pinion gears engage a further gear wheel connected to drivethe ground wheels. In one embodiment, the further gear wheel contains adifferential arrangement.

It has now been realised that plural power path arrangementsincorporating even torque sharing capability are of significant benefitto transfer gearing elsewhere in automotive transmissions, in particularbetween the main variable ratio gearbox and the axle (differential)drive, and also in other locations “up stream” of the axle differential.

In accordance with the invention there is provided an automotivetransmission comprising a transfer gear train for transmitting torquebetween an input rotatable member and an output shaft rotating aboutsubstantially parallel axes, the transfer gear train comprising an inputgear rotatable with the input member, an output gear rotatable with theoutput shaft, and a pair of intermediate gears each held simultaneouslyin mesh with the input gear and transmitting torque to the output gearto provide two power transmission paths, characterised in that theoutput shaft drives differential gearing arranged to distribute drivingtorque to a pair of ground engaging wheels. Preferably, one of the gearsin the transfer train is made movable in response to the transmittedtorque so as to even out power transmission between the two paths.However such torque sharing can also be achieved by other means, such asby controlling gear tooth flank position errors to within acceptably lowlimits.

The input gear is preferably made smaller than the output gear so thatthe transfer gear train provides a reduction ratio. The differentialgearing may therefore have a lower reduction ratio, even substantially1:1. This enables it to be made considerably smaller and lighter. Pitchline velocities in the transfer gearing and in the rest of thetransmission driven by it are also reduced, giving quieter operation.The sizes of the gears in the transfer gear train and differential canbe smaller than in a conventional single power path driveline, whichmuch reduces gear weight (roughly proportional to diameter squared) andvery much reduces gear moments of inertia (roughly proportional todiameter cubed).

In current driveline designs for passenger cars, the axle bevel gear hasa large diameter, typically providing a reduction ratio of about3:1—even larger for heavy vehicles. Any diminution in this reductionratio increases gear loadings “up stream”, including in the main gearboxand transfer gearbox (if present). Current driveline proportionstherefore represent a trade off between minimum gearbox size and maximumacceptable axle ratio. The plural power path transfer gear train of thepresent invention provides a high capacity, compact power transfer paththat can handle the higher torque loads arising from the use of lowratio axle differentials, and may itself be used to provide a reductionratio, thereby reducing torque loadings on the main gearbox.

A smaller axle differential gives a greater ground clearance, which isimportant in off-road vehicles. Incorporating speed reduction in thetransfer gear train also enables the overall reduction ratio of thetransmission to be maintained, whilst using a smaller reduction ratio atthe differential gearbox. This is beneficial in high performancevehicles such as racing cars, with engines operating at high RPM andwhich therefore require a high overall driveline reduction ratio. Forsuch applications, the transfer gear train of the present inventionagain gives lower pitch line velocities and smaller, lighter, quieter,lower inertia differential gearboxes and final drivelines.

The two power paths enable the transfer gear train to be made smallerand lighter than a conventional transfer gearbox of equivalent duty. Tofacilitate assembly of the various gears in proper meshing engagementand to provide flexibility in the available gear ratios and in thecentre spacing between the input and output members, a furtherintermediate gear is preferably provided in each power transmissionpath.

The torque responsive movement referred to above may for example be ofthe input gear. In one possible arrangement, the input shaft and therotational axes of the co-operating intermediate gears all liesubstantially in a common plane.

Where the various gears are spur gears, the input shaft may be made freeto move (e.g. pivot) very slightly away from this plane. Then, whenthere is flank to flank contact between the teeth of the input gear andonly one of the intermediate gears, the torque applied to the inputshaft and the reaction at the contacting tooth flanks will form a couplecausing the input shaft to move out of the common plane. This movementcontinues until there is flank to flank contact at the otherintermediate gear as well. Even torque sharing between the two powertransmission paths is therefore achieved.

Where the various gears are single helical gears, the input gear may befree to pivot about an axis normal to the common plane. Out-of-balanceforces acting in the direction of the input shaft axis and arising fromuneven torque sharing between the two power paths will rotate the inputgear about the pivot axis in a direction tending to reduce theout-of-balance forces, and hence evening up the torque sharing.Operation of such a mechanism is more fully explained in GB 1434928. Toreduce or substantially eliminate thrust loads on the gear shafts, thegears may be mounted to their shafts by helical splined connections, thehelix being of the same hand and having the same lead as the gear teeth.The splined connection may itself allow the input gear to pivot about anaxis passing through the input gear.

Similar arrangements are possible in which even torque sharing orcompensation is provided by movement of the output gear.

The torque compensating gear may also be free to move slightly both in adirection normal to the common plane and/or along the axis of the inputrotatable member, as described for example in relation to the outputgears 24, 34 in EP 0244263. Another arrangement permitting suchtranslational movements of a torque compensating gear is describedbelow.

The invention in its various aspects, and its further preferredfeatures, are described below with reference to illustrative embodimentsshown in the drawings, wherein:

FIG. 1 is a diagrammatic front view of a transfer gear train embodyingthe invention;

FIG. 2 is a diagrammatic sectional view of an automotive transmissionincorporating the transfer gear train of FIG. 1;

FIG. 3 is a diagram showing a mounting arrangement for a torquecompensating single helical gear that may be used in embodiments of theinvention;

FIG. 4 is a diagrammatic section of a further differential unit whichmay be used in a modification of the FIG. 2 transmission for four wheeldrive vehicles, and

FIG. 5 is a diagrammatic section of a planetary differential unit whichmay be used in a further modification of the FIG. 2 transmission.

Referring to FIG. 1, a transfer gear train 10 for an automotivetransmission, comprises an input rotatable member in the form of aninput shaft 12 connected to the output of a variable ratio main gearbox(not shown). Input shaft 12 carries an input gear 14 which meshessimultaneously with a pair of first intermediate gears 16 a, 16 bmounted on intermediate shafts 18 a, 18 b journalled in a gearbox casing(not shown). The intermediate gears. 16 a, 16 b mesh with respectivefurther intermediate gears 16 c, 16 d mounted on shafts 18 c, 18 d. Thefurther intermediate gears 16 b, 16 c mesh simultaneously with an outputgear 20. In the drawings, the circles illustrate the pitch lines of thevarious gears shown. For simplicity, the gear teeth are not illustrated.

As shown in FIG. 2, the output gear 20 is fixed to the input shaft 26 ofan axle differential unit 34, having output shafts 36 a, 36 b fordriving respective road wheels (not shown). The differential unit 34 maybe of any known conventional kind, for example incorporating sliplimiting or lockup means.

To provide even torque sharing between the parallel power pathsconstituted by the gears 16 a, 16 c on the one hand and 16 b, 16 d onthe other, the input shaft 12 is mounted so that its end carrying thegear 14 is movable slightly out of the plane containing the shafts 18 a,18 b. The bearings 30 may allow pivoting of the shaft 12 about an axisnormal to the page (as indicated by the arrows 32) and the shaft 12 mayincorporate a splined or other connection for this purpose. Such torquecompensating movement is suitable for use with spur gears. If insteadthe gears 14, 16 a, 16 b, 16 c, 16 d are single helical, gear 14 may bemounted to the shaft 12 in the manner described in GB 1434928 to providethe necessary torque compensating movement. Similarly, if the variousgears are double helical, mounted in “herringbone” configuration, torquecompensation of the resulting compound gear 14 can be as disclosed in EP0244263.

FIG. 3 shows a further torque sharing arrangement for use with singlehelical gears. The compensating movement is provided by mounting gear 14to shaft 12 via a torque transmitting sleeve 40. Gear 14 is formed as aring gear having internal helical splines engaging complementary splines42 on the sleeve 40. Internal helical splines on the sleeve in turnengage complementary splines 44 on the shaft 12. The splines 42, 44 areshort and slightly crowned, to allow pivoting of the axes of shaft 12,sleeve 40 and gear 14 relative to each other. The lead and hand of thesplines 42, 44 are equal to the lead and hand of the helical gear teeth46. This ensures that axial forces on the sleeve 40 and gear 14, arisingfrom the transmitted torque, balance out. With this arrangement, thegear 14 is not only free to pivot slightly out of the plane normal tothe shaft 12, but can also translate slightly, both axially and normalto the plane passing through the axes of shafts 18 a and 18 b. Thisprovides improved torque sharing with respect to the mechanism of GB1434928.

Although in the drawings the torque sharing or compensation mechanism isshown applied to the input gear 14, it could equally be applied to anoutput or other gear meshing simultaneously with a pair of furthergears, the shafts of all three gears being substantially co-planar. Forexample, the torque compensating movement may be of the output gear,particularly in transfer gear trains providing a step-up ratio.

As shown, the input gear 14 is smaller than the output gear 20. Thistherefore provides a reduction ratio. The reduction ratio of the finaldrive differential gearbox 34 may therefore be made smaller. A smallerbevel gear 38 may therefore be used. This results in a much slimmergearbox 34, further reducing the weights and amounts of materialsrequired, and improving the vehicle ground clearance. Large overalltransmission speed reduction ratios may also be achieved, which can beadvantageous in high performance vehicles such as Formula 1 Grand Prixracing cars. These currently use servo-operated, clutchless gearshifts,the gearbox having close ratios so as to eliminate the requirement forsynchronisers. A large reduction ratio is needed in the transmissionfinal drive. The main gearbox output shaft is close to the ground, sothat the drive must be taken upwards to the rear wheels. The arrangementshown in FIG. 2 is suitable for such use. (For clarity, FIG. 2 shows themain transmission elements in co-planar configuration. When used inracing car transmissions, the gear train 10 may extend upwardly, i.e.the transfer gear train 10 is rotated from the position shown, relativeto the axle differential 34, about the shaft 26, so as to place theshaft 12 at a lower level than the shaft 26).

A large centre spacing between the shafts 12, 26 is possible, whilstkeeping the size of the gears 16 a, 16 b, 16 c, 16 d reasonably small.The transfer gearbox is accordingly compact, lightweight, quiet,efficient, has low inertia, and is capable of handling high shaftpowers. Still further intermediate gears can be added into the powertransmission paths, as desired.

It is also possible to eliminate the further intermediate gears 18 c, 18d to produce a four gear arrangement in which the gears 18 a, 18 b meshdirectly with gear 20 (the centres of gears 18 a, 14 and 18 b stillbeing substantially in line to provide torque sharing). However, thegear ratios must then be carefully selected so as to achieve propermeshing of the gear teeth and so as to avoid clashing of gears 14 and20. Only a limited number of gear ratios and offsets between shafts 12and 26 are therefore available, which can be found e.g. by numericalmeans. Some examples are tabulated below.

Numbers of teeth Gears Gear 18a, Gear 14 18b 20 5 13 23 5 19 29 5 25 355 41 31 5 37 47 5 43 53 5 49 59 5 55 65 5 61 71 5 67 77 5 73 83 5 79 895 85 95 10 20 40 10 32 52 10 44 64 10 56 76 10 68 88 10 80 100 10 92 11210 104 124 10 116 136 10 128 148 10 140 160 10 152 172 10 164 184 10 176196 15 27 57 15 33 63 15 45 75 15 51 81 15 63 93 15 69 99 15 81 111 1587 117 15 99 129 15 105 135 15 117 147 15 123 153 15 135 165 15 141 17115 153 183 15 159 189 15 171 201 15 177 207 15 189 219 15 195 225 15 207237 15 213 243 15 225 255 15 231 261 15 243 273 15 249 279 15 261 291 15267 297 30 72 132 30 108 168 30 120 180 30 144 204 30 156 216 30 180 24030 192 252 30 216 276 30 228 288 30 252 312 30 264 324 30 288 348 30 300360 30 324 384 30 336 390 30 360 414 30 372 432 30 396 456 30 408 468 30432 492 30 444 504 30 468 528 30 480 540 30 504 564 30 516 576 30 84 144

FIG. 4 shows a further differential unit 24 for use in four wheel drive(or higher) transmissions. Instead of being fixed directly to the shaft26, the gear 20 is formed as a ring gear attached to a differentialcarrier 50 rotatable in a casing 52 of the differential unit 24. Theshaft 26 forms one of the output shafts of the differential unit 24. Afurther output shaft 28 takes the drive to a further pair of vehicleroad wheels (not shown), via a further axle differential unit (notshown), similar to unit 34. The transmission arrangement is otherwisesimilar to FIGS. 1 and 2, with the ring gear 20 engaged simultaneouslyby the gears 16 c, 16 d (only gear 16 c is visible in FIG. 5; in “fourgear” transfer train arrangements, ring gear 20 is of course engageddirectly by gears 16 a, 16 b). The differential unit 24 splits the inputtorque equally between the output shafts 26, 28. Like the axledifferential 34, it may incorporate conventional slip limiting andlock-up means. The transfer gear train 10 may be rotated about theshafts 26, 28, from the position shown, to any desired configuration.

FIG. 5 shows a modification of the FIG. 4 arrangement, using a planetarydifferential 60. The ring gear 20 has external teeth for engagement bythe gears 16 c, 16 d (or 16 a, 16 b) as in FIGS. 1, 2 and 4. It also hasinternal teeth engaged by a plurality of (e.g. three) planet gears 62,which in turn engage a sun gear 64. The gears 62 are journalled in aplanet carrier 66, which drives one output shaft 26. The sun gear 64drives the other output shaft 28. The planetary differential 60 providesan uneven torque split between the shafts 26: 28 in the ratio

d_(s)+d_(p)/ 2: d_(s),

where d_(s) is the diameter of the sun gear 64 and d_(p) is the diameterof the planet gear 62. The differential 60 may be provided withotherwise conventional lock-up or slip limiting means, acting betweenthe planet carrier 52 and sun gear 64.

What is claimed is:
 1. An automotive transmission comprising a transfergear train for transmitting torque between an input rotatable member andan output shaft rotating about substantially parallel axes, the transfergear train comprising an input gear rotatable with the input member, anoutput gear rotatable with the output shaft, and a pair of intermediategears each held simultaneously in mesh with the input gear andtransmitting torque to the output gear to provide two power transmissionpaths; wherein the output shaft drives differential gearing arranged todistribute driving torque to a pair of ground engaging wheels.
 2. Thetransmission of claim 1 wherein the input gear is smaller than theoutput gear.
 3. The transmission of claim 1 wherein one of the gears inthe transfer train is made movable in response to the transmitted torqueso as to even out power transmission between the two paths.
 4. Thetransmission of claim 3 wherein the torque responsive movement is of theinput gear.
 5. The transmission of claim 1 wherein a furtherintermediate gear is provided in each power transmission path.
 6. Thetransmission of claim 1 wherein the rotational axis of one of the gearsin the gear train which is movable in response to the transmitted torqueso as to even out power transmission between the two paths, and therotational axes of two other gears in the gear train which meshsimultaneously with the movable gear, all lie substantially in a commonplane.
 7. The transmission of claim 6 comprising spur gears wherein theshaft of the movable gear is free to move away from the common plane. 8.The transmission of claim 6 comprising single helical gears wherein themovable gear is free to pivot about an axis normal to the common plane.9. The transmission of claim 8 wherein the movable gear is mounted to ashaft by a helically splined connection, the helix of the splines beingof the same hand and having the same lead as the helix of the gearteeth.
 10. The transmission of claim 6 wherein the movable gear is freeto translate out of the common plane and/or along its rotational axis.11. The transmission of claim 10 wherein the movable gear is mounted tothe shaft via a torque transmitting sleeve, there being helicallysplined connections between the shaft and the sleeve and between thesleeve and the movable gear.
 12. The transmission of claim 1 wherein oneof said gears comprises two single helical gears mounted to a commonshaft by helical splined connections, the common shaft being coupled tothe input or output shaft by a drive connection that allows axial andradial displacement of the common shaft.